Hydrostatic thrust bearing system

ABSTRACT

A hydrostatic thrust bearing system characterized in that the hydrostatic bearings are axially movable and are held urged against the pressure plate in mutually opposite directions by hydraulic pressure created in a chamber associated with a respective bearing; and in that a servo valve is provided for co-action with an axial-position sensor for controlling the amount of hydraulic fluid fed to a respective chamber and also to control the hydraulic pressure acting on a respective bearing such as to take-up axial loads which vary in magnitude and direction.

FIELD OF INVENTION

The present invention relates to a hydrostatic thrust bearing system fora radially journalled axle or shaft provided with a pressure plate orcollar, the thrust bearing system having an external configuration whichenables the system to be mounted around the rotatable axle and includinga hydrostatic thrust bearing on each side of the pressure plate, and inwhich system each of the thrust bearings incorporates at least one oilpocket to which hydraulic fluid is fed from a hydraulic fluid source,preferably a constant flow pump.

STATE OF THE ART

Thrust bearings of this kind intended for taking up large axial forcesare to be found in many technical fields.

One technical field in which such thrust bearings are used is found inthe fibre beating mills of the paper industry, in which mills axialforces in the order of five tones or more can be generated in thebeating process, even in mills of modest dimensions, and which also varyduring the beating or pulping process.

For the purpose of maintaining a beating gap of predetermined width, thejournalling problems are normally solved by arranging the radialbearings of the shaft or axle in a bearing box or a bearing package andthen adjusting the whole of the bearing package axially with the aid ofhydraulic systems in dependence on the axial loads that occur, in orderto adjust the beating gap.

There is normally used in this regard a servo value which is in directcontact with the bearing package. The servo valve ensures that thebeating gap can be adjusted smoothly and continuously to a tolerance of1/100 mm while the beater is in operation, despite large variations inaxial loads.

Also known to the art is a combined hydrostatic/hydrodynamic bearingsystem for beating apparatus in paper manufacturing mills, in which thetwo end surfaces of a collar provided on a shaft or axle constitutepiston surfaces which work in a surrounding pressure chamber in a mannerto enable the collar, and therewith the shaft, to be moved axially inthe pressure chamber (see Wo 86/01434-Sunds Defibrator).

Variations on the axial load on the shaft are counteracted by means ofcompensatory fluid pressure on the end surfaces of the piston, such asto maintain a beater gap of predetermined width.

This publication, however, does not teach a hydrostatic bearing of thekind to which the present invention relates, i.e. a hydrostatic bearingwhich includes an oil pocket which exhibits an inner and an outer twoends annular sealing gap and has a constant leakage flow. In the case ofthis known bearing system, the radial gap between the chambers definedby the piston always has a constant width.

Various further types of hydrostatic bearings are known to the art. Oneessential characteristic of a hydrostatic bearing is that it isdependent on the supply of fluid from an external pressure source. Thisexternal pressure source ensures that metallic contact is avoided underall circumstances when the fluid supplied is evacuated through leakageor sealing gaps located in peripheral parts of the bearing.

Prior art hydrostatic bearings include so-called ring chamber bearingswhich are located opposite one another and which are mounted with apre-set gap width. Consequently, a great deal of precision is requiredwhen fitting the bearings.

DE-A-2 357 881 (Mannesmann-Meer AG) teaches a hydrodynamic thrustbearing provided with a flanged pressure plate against which a pluralityof slide shoes are urged by means of hydraulic pistons. These slideshoes move over an oil film having a thickness of between 0.008 and0.012 mm. This oil film is created in a manner similar to that employedin other well known slide bearings, i.e. the film is built up betweenthe two bearing surfaces as a result of the relative movementtherebetween. The thickness of the oil film is therewith contingent onthe speed at which the shaft or axle rotates and has no lubricating orcarrying function when the axle is stationary.

In the case of these known hydrodynamic bearing systems the separatehydraulic pistons are intended to take up the load uniformly around therotationally symmetrical pressure plate, so as to equalize pressuresurges and to provide a "play free" arrangement.

Other types of known hydrostatic thrust bearing systems which relate tothe bearing system of the present invention, i.e. bearing systems whichinclude oil pockets, e.g. annular hydraulic chambers, are founddescribed and illustrated in SE-B-8105404-1 (Svenska Rotormaskiner),F16C21/00, DE-A-23 57 881 (Mannesmann-Meer) and DE-A-24 48 785(Kugelfischer Georg Schafer).

OBJECTS OF INVENTION

An object of the present invention is to provide a simplifiedhydrostatic thrust bearing system of the aforesaid kind which willafford improved precision in the bearing under conditions of radicallyvarying axial loads, despite the simplified construction of the bearingsystem, and thereby enable the shaft or axle to be rotated at higherspeeds.

SUMMARY OF THE INVENTION

A hydrostatic thrust bearing system according to the invention whichfulfils this and other objects is of the kind set forth in theintroduction and is mainly characterized in that the hydrostaticbearings are axially movable and are held urged against the pressureplate in mutually opposite directions by means of hydraulic pressurecreated in a chamber associated with a respective bearing; and in thatmeans, e.g. a servo valve, is provided for co-action with anaxle-position sensing means for controlling the amount of hydraulicfluid fed to a respective chamber and also to control the hydraulicpressure acting on a respective bearing such as to take-up axle loadswhich vary in magnitude and direction.

The hydrostatic bearing according to the invention is axially movableand is held urged against the pressure plate by means of hydraulicpressure acting in the chambers of respective bearings in mutuallyopposite directions. Control means, e.g. in the form of a servo valve,is intended to co-act with a sensor which senses the axial position ofthe axle or shaft, such as to control the amount of hydraulic fluidwhich is supplied to respective chambers and to control the hydraulicpressure acting on respective bearings, thereby taking up axle loadswhich vary in magnitude and direction.

With a bearing system constructed in accordance with the invention, theaxial position of the axle is adjusted smoothly and continuouslyirrespective of the load on the axle and the speed at which it rotates.

As opposed to known bearings, e.g. the bearing system taught by theabove mentioned German published specification 2 357 881, which do notafford smooth, continuous adjustment of the axle or shaft in response tovarying axial loads, but merely afford a limit position function, abearing system constructed in accordance with the invention enablespressure and the fluid supplied to the two hydraulic chambers to bedistributed in a smooth and continuous fashion.

There is preferably used to this end a hydraulic servo valve, e.g. acopying valve, which will distribute pressure and fluid quantitycontinuously and smoothly between the two chambers.

Because the invention bearing system has two axially movable hydrostaticbearings which are urged hydraulically against the pressure plate inmutually opposite directions, it is not necessary to move the entirebearing box or bearing package axially, thereby enabling the bearings tobe of simpler construction and improving the reliability of the bearing.

The width of the generated fluid-throttling gap is a function of theload, i.e. the fluid pressure in accordance with a known formula, suchthat a greater fluid flow will result in a larger gap, while a greaterload results in a higher fluid pressure, which in turn results in asmaller gap. A higher viscosity of the fluid will also result in anincrease in gap width.

The invention thus insures that the gap width of the two leakage orsealing gaps will be adjusted automatically, i.e. the gap isself-adjusting.

A preferred embodiment of the invention in which each hydraulic fluidsource produces a constant hydraulic flow which generates a leakage orsealing gap which departs from respective oil pockets and which isoperative in creating a throttling effect and thereby pressurizing thefluid, is characterized in that said control means, e.g. the servovalve, when the hydrostatic bearings are subjected to varying axialloads, is intended to maintain a hydraulic pressure which variesproportionally to the load, such that the leakage or sealing gaps obtaina self-adjusting gap width which varies in inverse proportion to thehydraulic pressure.

The two thrust bearings thus work in the manner of axially actinghydraulic pistons which, in accordance with a further embodiment of theinvention, are each guided radially in a respective cylinder e.g.through the intermediary of a Teflon band placed around the innerperipheral surface between sealing rings. This will avoid metal-on-metalcontact between the bearings and the inner surfaces of the cylinders.

Due to the rotational symmetry of the bearings and the cylinder, it isnormally sufficient to guide the bearings along one of the surfaces.

The two hydrostatic bearings may have oil pockets of varyingconfiguration and in varying numbers. It is preferred, however, that thebearings have an outer configuration which will enable the bearings tobe fitted around the rotatable shaft.

In practice, the hydrostatic bearings will have a so-called ring-chamberconfiguration, i.e. the oil pocket is annular and includes radiallyinner and outer annular leakage or sealing gap.

The supply of hydraulic medium, e.g. oil, to the hydrostatic bearings ispreferably effected with the aid of constant flow pumps, although othermore exclusive means may be used to this end.

BRIEF DESCRIPTION OF THE DRAWINGS

The invention will now be described in more detail with reference toexemplifying embodiments thereof and also with reference to theaccompanying schematic drawing.

FIG. 1 is a vertical section through part of an axle which is providedwith a pressure plate and which incorporates the inventive hydrostaticthrust bearing system, said figure illustrating the principle inaccordance with which the inventive system functions.

FIG. 2 is a vertical section of a working axle which incorporates theinventive bearing system and which forms part of a working machine, saidfigure also showing radial bearings on respective sides of the thrustbearing system.

FIG. 3 is a vertical section through a somewhat modified embodiment.

In the Figures like components have been identified with likereferences.

DESCRIPTION OF PREFERRED EMBODIMENTS

In FIG. 1 the reference 1 designates an axle which is assumed to besubjected to axial loads of varying magnitude and direction. In order toenable these loads to be taken up, a ring or oil bearing 2 and 3 is heldurged hydrostatically against a respective side surface of the pressureplate or collar 1a, these bearings 2, 3 having an external configurationwhich enables them to be mounted around the axle.

Each bearing is provided with a peripherally extending oil pocket 2a, 3ato which oil is supplied from a respective constant flow pump 4 and 5driven by a motor 7. Inner and outer ring leakage or sealing gaps extendfrom the oil pockets 2a, 3a. The oil leaking from the pocket 2a throughthe inner gap is conducted parallel with and circumferentially aroundthe axle 1, and departs from the bearing in a flow A. The flow of oilleaking through the outer leakage gap is referenced D.

Corresponding leakage flows from the oil pocket 3a through respectiveinner and outer leakage gaps are referenced H and E. The constant flowpumps 4, 5 which supply the oil pockets 2a, 3a with oil are referenced 4and 5, respectively. The pressure lines 4a and 5a extending fromrespective pumps 4, 5 are connected to the ring oil bearing at C and F,respectively.

Each of the bearings 2, 3 is arranged for axial movement in a respectivecylinder 12, 13 and is also guided radially therein. In the illustratedembodiment this radial guidance of the bearings is achieved by means ofa respective Teflon band 2b, 3b located on the inner peripheral surfacebetween sealing rings 2c and 3c. Further sealing rings 2d, 3d areprovided on the outer peripheral surface of respective bearings.

The bearings 2, 3 in respective cylinders 12, 13 are subjected tohydraulic pressures which, when the axial load varies or changes,results in automatic adjustment, or self-adjustment, of the width of theleakage or sealing gaps extending from respective oil pockets 2a, 3a.

Each of the bearings 2, 3 therewith functions as an axially workinghydraulic piston, which in the illustrated embodiment is controlled by aservo valve 8. This valve is of the copying-valve kind and has a slidewhich is in physical contact with the axle 1 via separate means, suchthat the pressure in the hydraulic cylinders 12, 13 is adjusted inresponse to the slightest change in the axial position of the axle 1.

The copying valve 8 is maintained under constant oil pressure, e.g. apressure of 70 bars, by means of a pump 6. The valve pressure ismaintained constant by means of a pressure limiting valve 9.

The copying valve 8 is of a conventional kind and is intended to insurethat the space behind respective bearings 2, 3 in the cylinders 12, 13stands under the prevailing pressure, via lines 8a and 9a withconnections B and G, respectively.

In the illustrated embodiment the axial position of the axle 1 isadjusted by changing the setting of the valve 8 directly, with the aidof a remote-controlled stepping motor 10. Although not shown, anadjustment screw may be used as an alternative to the motor 10. Thestepping motor 10 operates in co-action with a sensor 10b, through aline 10a, and receives from the sensor information concerning axialdisplacement movements of the axle.

When the sensor 10b detects axial movement of the axle 1, for examplemovement of the axle to the left in the drawing, a corresponding signalis sent to the stepping motor 10 which in response thereto displaces theslide 8c of the copying valve 8 so as to adjust the valve to one of thethree symbolically illustrated valve positions. Oil under pressure istherewith passed from the pump 6, through the line 8a and the connectionB, to the chamber of the cylinder 12 accommodating the bearing 2,whereupon the bearing will move the axle to the right in the drawing,i.e. a self-adjusting displacement operation is carried out, which isregistered by the sensor 10b co-acting with the stepping motor 10, whichin response to the signal sent by the sensor 10b switches the valve tothe valve position corresponding to the central position of the threesymbolically illustrated valve positions.

Should the axle 1 be moved to the right in the figure, the steppingmotor will instead switch the valve 8 to a valve position correspondingto the left-hand valve position of the illustrated valve positions. Thiswill then result in an increase in pressure in the chamber of the rightcylinder 13.

FIG. 2 illustrates a working example in which the axle 1 forms part of amill for beating fibre stock in the paper industry. Such beating millaxles are subjected to large axial loads, e.g. loads in the order of 5tons or more, and may have a diameter in the region of 300 mm in thevicinity of the bearings 2, 3.

Arranged on each side of respective thrust bearings 2, 3 are stationaryradial bearings 15, 16, which in the illustrated embodiment compriseconventional cylindrical roller bearings. Bearings of this kind permitaxial displacement of the axle 1 through a distance of ±10 mm.

The thrust bearing system of the FIG. 2 embodiment is in principle ofsimilar construction to the system of the FIG. 1 embodiment.

FIG. 2 also illustrates the four connections B, C, F and G for oil underpressure, these connections being identified by the same references. Theleakage flows A, D, E and H departing from the inner and outer leakageor sealing gaps of respective bearings are received and conducted awayvia connections which are identified by references corresponding tothose used in FIG. 1.

The axle 1 and associated radial and thrust bearings 15, 16 and 2, 3 arehoused in a housing or casing 17.

The bearing system of FIG. 2 provides a high degree of precision,despite rotation of the axle and the varying axial loads to which it issubjected. In comparison with conventional axial roller bearings ofcorresponding size, which can be normally driven at a speed of at most800/1000 rpm, with forced lubrication, possibly up to 1500 rpm, abearing constructed in accordance with the invention enables the axle orshaft to be driven at twice the speed, i.e. at a speed of about 3000rpm.

The improved position afforded by the inventive bearing system is mainlydue to the aforedescribed direct servo steering of the hydrostaticbearings, i.e. it is not necessary to steer the entire bearing housingor bearing box, such bearing housings or bearing boxes of conventionalsystems normally incorporating radial bearings.

It follows from what has been stated above that the hydraulic systemfrom a position of hydraulic equilibrium in which there is no axialthrust on the axle 1 shall continuously increase or decrease,respectively, the pressures in the opposite hydraulic chambers when anouter axial thrust is developed in any direction. Further, the axialposition of the axle shall be changed by varying the volumes ofhydraulic fluid in the two opposite chambers.

A device having this capability is a hydraulic servo device of thegeneral kind described above with reference to valve 8, stepping motor10 and sensor 10b.

The said valve 8 is arranged when feeded with a constant system pressureP, at the position of hydraulic equilibrium, to deliver half thepressure, i.e. P/2 to each connection or port B and G, respectively.

By offsetting the slide of the valve the pressures in B and G willincrease or decrease, respectively, in proportion to the displacement ofthe slide.

The slide is, as shown, via stepping motor 10 and sensor 10b (or anyother suitable means, for instance an electric device) re-coupled to theaxle 1. Then a displacement of the axle will be transformed to aproportional pressure increase/decrease in the respective chambers.

In addition, the preset position of equilibrium of the servo device mayby outside influence be changed so that a corresponding permanentdisplacement of the axle 1 is accomplished. This will take place whenvalve 8 permits a bigger or smaller, respectively, volume of hydraulicfluid to be present in the two opposite hydraulic chambers.

In practice this can be obtained by means of a manually operableset-screw or a remote-controlled step motor.

In a practical example the system pressure is 70 bar. The hydraulicchambers in communication with connections or ports B and G,respectively, have each a pressure-exposed area of 1000 cm².

In unloaded condition the pressure in each chamber will be P/2=35 barresulting in a bearing thrust of 35·1000=35 tons.

If then an axial thrust is developed the pressure will decrease/increasein the opposite chambers which means that the thrust on one of thebearings will decrease as much as it is increased on the other one.

This can be illustrated in a diagram where pressure and resultingbearing thrust of each chamber is given as a function of the outergenerated axial thrust.

There are, however, in practice other ways of controlling said process.Thus, by using a micro-processor it is for instance possible to simulatethe function of the servo valve.

A micro-processor will then be feeded with data giving information onthe axial position of the main axle--several times per second--and thisinformation is processed into one or more output signals.

These signals may then by used via hydraulic throttle valves toindividually control pressures and fluid volumes in respective chambers.A device of this kind will also fall within the scope of the attachedclaims.

In the FIG. 3 embodiment the servo steering system is of the generalkind described above with reference to FIG. 1. Corresponding somewhatdifferent parts have been provided with a prime sign (').

The main difference relative to the previously described embodimentsresides in that each bearing cylinder 12', 13' does not completelysurround a piston-like thrust bearing such as the ones denoted 2 and 3,respectively, in FIGS. 1 and 2.

Instead each thrust bearing 2', 3' has a stepwise outer peripheralconfiguration, and the cylinders 12', 13' a corresponding matingstepwise inner peripheral configuration to coact with each bearing. Onlyone stepwise surface of each cylinder 12', 13' and thrust bearing 2',3', respectively, have then to be carefully machined.

The general function will be the same but the manufacture of the bearingshown in FIG. 3 is substantially simplified.

INDUSTRIAL APPLICATION

A number of important advantages are gained by the inventive combinationof mutually opposing hydrostatic thrust bearings, hydraulic pressurecontact of the bearings with the pressure plate, and the aforesaidsmooth and continuous distribution of pressure and hydraulic fluid.

In addition to those particular advantages which are afforded when usinga hydrostatic bearing system which enables loads acting in variousdirection to be taken up and which is insensitive to load and axlespeed, and which enables the gap width to be adjusted automatically(self-adjusting), the bearing system according to the invention alsoenables the axial position of the axle or shaft to be adjusted smoothlyand continuously to a very high degree of accuracy, e.g. to an accuracyof 1/100 mm.

Other important advantages afforded by a hydrostatic thrust bearingsystem according to the invention are:

the possibility of driving the shaft or axle at considerably higherspeeds than with known bearing systems,

greater flexibility with regard to the loads that can be taken up,

longer useful life,

higher setting precision,

simpler and less expensive manufacture, and

reduced sensitivity to vibrations.

A bearing system according to the invention lends itself to furtherdevelopment and can be used in many different technical fields. The twohydrostatic bearings may be provided with any desired number of oilpockets, which may have any desired configuration. It will also beunderstood that the hydraulic medium used in the cylinders accommodatingthe hydrostatic bearings need not necessarily be oil and that anysuitable hydraulic medium may be used.

I claim:
 1. A hydrostatic thrust bearing system for a radiallyjournalled axle or shaft (1) provided with a pressure plate or collar(1a), the thrust bearing system having an external configuration whichenables the system to be mounted around the rotatable axle (1) andincluding a hydrostatic thrust bearing (2, 3; 2'a, 3'a) on each side ofthe pressure plate (1a), and in which system each of the thrust bearingsincorporates at least one oil pocket (2a, 3a; 2'a, 3'a), to whichhydraulic fluid is fed from a hydraulic fluid source, characterized inthat the hydrostatic bearings (2, 3; 2', 3') are axially movable and areheld urged against the pressure plate (1a) in mutually oppositedirections by means of hydraulic pressure created in a chamberassociated with a respective bearing; and in that means, is provided forcoaction with an axial-position sensing means (10b) for controlling theamount of hydraulic fluid fed to a respective chamber and also tocontrol the hydraulic pressure acting on a respective bearing such as totake-up axial loads which vary in magnitude and direction.
 2. A systemaccording to claim 1, in which each hydraulic fluid source (4, 5)produces a constant hydraulic flow which generates a leakage or sealinggap which departs from respective oil pockets (2a, 3a; 2'a, 3'a) andwhich is operative in creating a throttling effect and therebypressurizing the fluid, characterized in that said control means, whenthe hydrostatic bearings (2, 3; 2', 3') are subjected to varying axialloads, is intended to maintain a hydraulic pressure which variesproportionally to the load, such that the leakage or sealing gaps obtaina self-adjusting gap width which varies in inverse proportion to thehydraulic pressure.
 3. A system according to claim 1, characterized inthat each thrust bearing (2, 3; 2', 3') is mounted for axial movement inthe chamber of a respective hydraulic cylinder (12; 13), such that thehydraulic pressure acts against the end surface facing the oil pocket ofthe bearing.
 4. A system according to claim 3, characterized in thateach bearing is radially guided in respective cylinders (12, 13; 12',13'), e.g. by means of a Teflon band (2b; 3b) located on the innerperipheral surface between sealing rings (2c, 3c).
 5. A system accordingto claim 4, characterized in that each axial bearing (2, 3; 2', 3') isprovided with peripherally extending outer sealing rings (2d, 3d; 2'd,3'd) on each side of a peripheral, outer flared part of one channel torespective oil pockets (2a, 3a; 2'a, 3'a).
 6. A system according toclaim 1, characterized in that said control means includes a servo valvein the form of a copying valve (8) which has a slide (8c) that can beadjusted in dependence on the axial position of the axle, e.g. by meansof a stepping motor
 10. 7. A system according to claim 1, characterizedin that the outer peripheral surface of each thrust bearing (2', 3') hasa stepwise configuration coacting with a mating stepwise innerperipheral surface on each cylinder (12', 13').
 8. A system according toclaim 1, characterized in that the means for controlling the amount ofhydraulic fluid fed to a respective chamber includes a micro-processor.9. A system according to claim 1, wherein said hydraulic fluid sourceincludes a constant flow pump (4, 5).
 10. A system according to claim 1,wherein said means for controlling the amount of hydraulic fluidcomprises a servo valve (8).